This invention relates generally to fluid cooled bearings for rotating members, and more particularly to an improved fluid cooling chamber in the bearing housing for such bearings wherein the bearings preferably have self contained lubrication systems and to an improved closed type circulating system operatively associated with the improved fluid cooling chamber all operative for maintaining the bearing below a predetermined maximum temperature.
The use of a chamber for cooling fluid to maintain bearings and/or the lubricant used with such bearings from exceeding a predetermined maximum temperature is known to the prior art as is shown in U.S. Pat. Nos. 3,806,210; 2,098,683; 2,344,240; 2,249,021 and 2,238,925.
The present invention is particularly applicable to water cooled journal bearings for single-stage steam turbines which are supplied lubricant by self contained lubrication systems having at least one oil slinger ring rotated by the turbine shafts which are rotatably mounted in such water cooled journal bearings.
These water cooled journal bearings are generally mounted in bearing housings and for reasons of proper alignment and concentricity of the rotating and stationary parts of the steam turbine, the bearing housings are solidly connected or affixed to the steam turbines in any suitable manner as will be clear to those skilled in the art.
The bearing housings for such water cooled bearings include the oil sump for the self-contained lubricating systems through which at least one oil slinger ring is rotated on rotation of the shaft to deliver oil from the sump to the journal bearing, and an integral cooling system which includes, a cooling fluid flow chamber having a common wall with the oil sump and through which cooling water from any suitable source such as river water, cooling pond water etc. is passed and returned so that the common wall serves as an elementary heat exchanger in the bearing housing.
In general, fluid cooling systems for such water cooled journal bearings are designed to achieve maximum heat transfer rates across such common wall with the oil sump for the self contained lubrication system because this will result in a relatively lower bearing temperature. Lower bearing temperatures allow for the selection of less costly bearing materials and promote maximum bearing life because optimum oil film thickness can be maintained between the journal bearings and the shafts rotatably journaled therein at such lower temperatures as will be more fully explained below.
In the selection of less costly bearing materials it is recognized by those skilled in the art that the use of babbitt lined bearing shells are both technically and economically desirable. However, inherent with babbitt lined bearing shells is that deformation thereof cannot be avoided where the bearing temperatures exceed 250.degree. F. Therefore in all fluid cooled babbitt type bearings, the bearing temperatures for all operating conditions must be maintained below this critical limit.
In order to understand the factors which influence the bearing temperature of water cooled bearings having self contained lubrication systems, it is essential to recognize the actual sources of heat flow to or from the journal bearings and the basic principles of heat transfer involved which affect such heat flow.
Additionally, the effect that increased temperature has on the pumping capacity of the oil slinger rings in the self contained lubrication system for the journal bearing must also be taken into account as is more fully set forth below.
Thus heat at the journal bearing comes from three sources. First, by conductivity from the hot turbine casing to the bearing housing, second by conductivity along the shaft from the hot temperature region in the turbine to the journal bearing, and third due to viscous shear of the oil film which separates the journal bearing from the shaft when the turbine shaft is rotating.
Treating first with the equation which relates the variables in respect of heat transfer or heat flow due to conductivity, reference may be had to any well known heat transfer texts such as Introduction to Heat Transfer by Brown and Marco, 3rd. edition where such equation is generally set forth as follows: EQU 1. Q= K.times. .DELTA.T.times. A/L
where
Q= Heat flow rate- BTU/hour PA1 K= Conductivity coefficient- BTU/hr/in.sup. 2 /.degree. F. PA1 .DELTA. t= temperature Difference- .degree. F. PA1 a= average Section areas- inch.sup.2 PA1 L= Length applicable to .DELTA.T - inches PA1 k= thermal conductivity of Water BTU/hr/ft/.degree. F./ft PA1 D= equivalent hydraulic diameter- ft PA1 V= water velocity ft/hr PA1 .rho. = water density lb/ft.sup.3 PA1 u= water absolute viscosity- PA1 C.sub.p = water specific heat- BTU/lb/.degree.F. PA1 1. Support structure designed to minimize heat flow from the hot turbine casing. PA1 3. passage means which form a flow path for returning oil, the result of which is that a large part of the oil supplied to the bearing will be cooled while in the bearing and while returning from the bearing. PA1 1. It could avoid ecological restrictions due to Federal, State or Local ordinances. PA1 2. It would avoid the maintenance complications and installation costs where protection from freezing weather would be required with open systems, in that closed systems generally utilize a mixture of water and ethylene glycol as the coolant, and PA1 3. It would avoid restriction on operation when a draught occurs.
In this equation, if we assume iron or steel is used in the turbine casing and bearing housing, then K and .DELTA.T will be fixed by economic or service conditions and it is only possible to influence Q by a judicious selection of the ratio of A the average section area to L, the length applicable to .DELTA.T.
Thus, this equation teaches that minimum heat flow from the turbine casing to the bearing housing can be achieved by minimizing the contacting area and the cross sectional area of the joint between the operatively associated supporting structure of the turbine casing and the bearing housing and by making the length, distance or spacing between the turbine casing and bearing housing as large as technical and economic considerations will permit.
Conversely for maximum heat transfer from the journal bearing shell, the metal wall of the journal bearing supporting the bearing shell and disposed between the bearing shell and the fluid cooling chamber in the bearing housing should have as large an area as possible while the thickness of this metal wall should be minimized so that this wall can be as thin as good foundry practice will permit.
In the prior art devices analysis shows that the cross-sectional area of the structure rigidly connecting the turbine casing and the journal bearing housing is quite large and the distance between the turbine casing and the journal bearing casing relatively short. This construction in accordance with the above equation produces large conductive heat flow rates from the turbine casing to the interior of the bearing housing where the bearing shell is located.
Additionally, in the prior art devices, the thickness of the metal of the journal bearing shell to the cooling water chamber is relatively large or longer than necessary and the area of this support structure is frequently quite small.
Thus, the connecting structure for the bearing housing and the bearing housing design of the prior art devices have not been planned either to minimize conductive heat flow to the bearing or to maximize conductive heat flow from the bearing to the cooling fluid in the fluid cooling chamber in the bearing housing. Consequently, in such prior art bearing housing designs, heat removal from the bearing is largely dependent upon the quantity and temperature of the oil supplied to the bearing by the slinger rings which rotate with the shaft.
It has been proven through tests by various investigators that the quantity of oil delivered by oil slinger rings is highly dependent upon the viscosity of the oil in the sump. As the viscosity decreases, oil delivery by the rotating ring drops very quickly. Since the rate of heat generation due to viscous shear rises rapidly with speed on journal bearings, higher oil temperatures which cause a reduction in oil viscosity will occur at a time when the oil delivery rate from the slinger ring declines sharply. Bearing failure will occur when the quantity and viscosity of oil supplied are insufficient to maintain an adequate oil film thickness between the journal and bearing.
Therefore, with such prior art bearings, the operating limitations of the slinger rings will usually determine the maximum operating speed and temperature limit for the turbine.
The complex, inter-dependent relationship of oil slinger ring pumping capacity to shaft speed and to conducted temperature from the turbine casing is the current principal limitation to extended usage of the low cost, oil slinger ring lubrication system.
In prior art bearing housings, a relatively large cooling water chamber is commonly located below and adjacent to the oil sump, the intended purpose being to cool the oil therein as much as possible by providing a common separating wall of relatively large area. This would appear logical in view of the dependence of prior art bearings upon oil slinger ring pump capacity as explained above.
However, an analysis of such constructions, following modern heat transfer laws for forced convection will prove that this elementary heat exchanger is very inefficient, primarily due to the poor heat conductivity of oil and the low through flow velocities of both the oil in the sump and the cooling water through the fluid cooling chamber in the bearing housing.
Collateral to heat flow due to conductivity those, skilled in the art will recognize that during operation, there will develop a thin but stagnant film of water or other cooling fluid which adheres to the interior surface of the cooling fluid chamber. This film adversely affects the heat transfer rate from the bearing shell and the wall supporting the bearing shell to the cooling water passing through the cooling fluid chamber in the bearing housing or from the common wall between the sump and the fluid cooling chamber to the cooling fluid passing through the fluid cooling housing.
The existence of the thin but stagnant film of water on the interior surface of the fluid cooling chamber in the bearing housing is described in considerable detail in most modern texts on heat transfer.
The equation which relates the variables affecting the forced convection heat transfer across the film to maintain minimum temperature difference between the cooling water and the adjacent metal wall can be generally derived from the Brown and Marco Text above cited at Page 133 as follows: ##EQU1## where h.sub.c = film, heat transfer coefficient- BTU/hr/ft.sup. 2 /.degree. F.
Since water is the usual coolant in such water cooled journal bearings, the only variables which can be adjusted are V, the water velocity and D the equivalent hydraulic diameter and thus this equation can be reduced to the following: EQU 2. h.sub.c .about.V.sup.0.8 /D.sup.0.2
thus, if V the water velocity is large and D the equivalent hydraulic diameter is small, a high heat transfer rate will be achieved because the thickness of the stagnant film on the interior surface of the cooling fluid chamber will be small.
By applying the above principles of conductive heat flow and force convection heat transfer to the present invention the bearing and the oil temperatures have been so significantly reduced as to permit turbines utilizing improved water cooled bearings with self contained lubrication systems of the oil slinger ring type in accordance with the present invention to be operated at higher shaft speeds. The advantage of such higher shaft speeds is to generally improve turbine efficiency and these higher shaft speeds become possible from a design which provides improved pumping capacity of the oil slinger ring in the self contained lubrication system.
Broadly, therefore the principle characteristics of the improved water cooled bearing with a self-contained lubrication system in accordance with the present invention will have at least the following major features.
2. A cooling fluid chamber which, by reason of its location, geometry, and interior design, achieves improved direct cooling of the bearing shell so as to minimize dependence upon the cooling effect of oil supplied to the bearing by the slinger ring, achieves improved direct cooling of the oil to be pumped from the adjacent oil sump, and additionally intercepts heat flowing from the turbine casing to the oil sump, and
Further, water cooled bearing housings of the prior art type and those in accordance with the present invention are usually horizontally split. To simplify inspection or replacement of the bearing shell, the cooling jacket or chamber with connections thereto is invariably located in the lower or bottom half of the bearing housing.
As will be understood by those skilled in the art, it is desirable to maintain the entire upper and lower housing in which the bearing shell is journalled, and which therefore surrounds the bearing shell, as close to an isothermal condition as possible. However, with horizontally split bearing housings, some portions of the upper half of the housing cannot be in immediate proximity to the cooling fluid chamber in the lower half of the bearing housing. Large temperature differences are avoided in the cooling and lubricating arrangement in accordance with the present invention by making thick walled sections on that part of upper half of the bearing housing which is furthest removed from the cooling fluid chamber in the lower half of the bearing housing. This construction being dictated by the same principles defined in the conductivity equation above. Thus, where physical necessity requires a longer length, the section area must be increased to maintain the minimal metal temperature about the upper half of the bearing shell.
It is also essential to note that journal bearings having self-contained lubrication systems which are associated with steam turbines may be subjected to excessive bearing temperatures when the turbine is brought to a stop. This occurs when the turbine is brought to a stop first because delivery of the cool oil supply to the bearing ceases and second due to the fact that it may require as much as an hour or more to dissipate the high temperature heat stored in the turbine casing and rotor. In effect therefore, when an operating steam turbine is brought to a stop the first and second sources of heat flow as above outlined will continue.
In prior art bearing housings heat removal by direct conductance from the bearing shell to the cooling water chamber is relatively small. As a consequence of this, maximum permissable bearing temperature are largely dependent upon the quantity and temperature of the oil delivered by the oil slinger ring which rotates with the shaft. Therefore, at idling speeds or when the turbine is brought to a stop, bearing temperatures can and have become excessive, particularly where the steam temperatures at which the steam turbine operates is high.
Accordingly, to prevent the bearing temperature from becoming excessive after a steam turbine is brought to a stop, a reliable bearing system for such steam turbine must not only be designed for adequate heat removal capacity during normal operating conditions but additionally for dissipation of the "heat soak" effect which occurs after a hot turbine is brought to a stop.
This is accomplished in the present invention by achieving large heat removal capacity by direct conduction from the bearing shell to the cooling fluid chamber which is independent of turbine shaft speeds. Therefore, excessive bearing temperatures above 250.degree. F. are avoided even when the shaft is stationary.
In the bearing in accordance with the present invention, when the shaft is rotating, additional cooling is obtained from the cool oil supplied by the oil slinger ring. Therefore, even at maximum shaft speeds, the bearing temperature will not exceed the usually specified limit of 180.degree. F.
The overall heat exchange rate of the bearing housing design of the present invention is much greater than that of comparable prior art bearing housings both when the turbine is running and when it is stopped. Therefore this bearing in accordance with the present invention not only meets and solves the problems regarding excessive bearing temperatures when the turbine is idling or brought to a stop but further permits accomodation of the turbine not only to the open cooling fluid circulating system with which comparable prior art devices have operated as is well known and understood by those skilled in the art but additionally to cooling fluid circulating systems of the closed type which are now in growing demand.
The conventional open system is one where water is piped to the turbine bearing housing from a lake or river and the heated effluent water is piped away to a sewer system. In a closed system, the coolant is circulated from a reservoir to and through the bearing housings and then the heated effluent coolant is passed to a heat exchanger, the cooled effluent coolant then being returned to the reservoir for recirculation to the bearing. Such closed systems utilizing a fan and radiator type heat exchanger are commonly found on automobile engines and on portable compressors.
For many applications a closed cooling fluid circulating system would be desirable for at least the following reasons:
Water Cooled Bearings in accordance with the prior art have never been successfully applied to closed fluid cooling systems except in isolated instances where both the inlet steam temperature and operating speeds of the turbine were low.
However, bearings in accordance with the present invention can be applied satisfactorily to either open or closed systems for steam inlet temperatures to 750.degree. F. and speeds to at least 5000 RPM.
In the cooling fluid circulating system of the closed type which is hereinafter illustrated for use with the improved cooling and lubrication arrangement for bearings in accordance with the present invention, there is further illustrated a "totally closed system concept" i.e. one in which the bearings of an auxiliary device can also be cooled by the same closed type fluid cooling system associated and used with the improved cooling and lubrication arrangement for bearings in accordance with the present invention.